Cooler distributor for a heat exchanger

ABSTRACT

A distributor ( 12 ) for an evaporator ( 22 ) in a vapor-compression circuit ( 10 ) comprises an elongate body ( 50 A.  50 B) defining an inlet portion ( 56 A) and first and second distal ends ( 58 A,  58 B). The body ( 50 A) being positionable along an interior sidewall ( 42 ) of the evaporator ( 22 ). The inlet portion ( 56 A) positionable opposite an inlet port ( 38 ) of the evaporator ( 22 ) such that refrigerant entering the evaporator ( 22 ) engages the inlet portion ( 56 A) of the distributor ( 12 ). The first and second distal ends ( 58 A,  58 B) extending outwardly towards opposing side ends ( 34 A,  34 B) of the evaporator ( 22 ). The elongate body ( 50 A,  50 B) cooperates with the sidewall ( 42 ) of the evaporator ( 22 ) to define a channel ( 49 ) having a generally uniform cross-sectional area extending from the first distal end ( 58 A) to the second distal end ( 58 B) such that refrigerant entering the channel ( 49 ) near the inlet portion ( 56 A) is initially substantially contained between the sidewall ( 42 ) of the evaporator ( 22 ) and the elongate body ( 50 A,  50 B). The elongate body ( 50 A,  50 B) defining discharge ports ( 60 A,  60 B) to disperse refrigerant from the channel ( 49 ) into the evaporator ( 22 ).

BACKGROUND

Typical refrigeration and air conditioning systems rely on vapor-compression cycles to transfer heat from one location to another for the purposes of cooling or heating an enclosed space. Such vapor-compression cycles comprise a compressor, a condenser, an expansion device and an evaporator connected to form a closed-loop circuit. In “chiller” systems, the vapor-compression circuit is used to facilitate cooling of multiple spaces within a building. Each component of the system is connected by a length of piping that conducts a working fluid, such as a chiller refrigerant, through the circuit. The compressor controls the flow of the chiller refrigerant through the circuit to adjust the amount of temperature control that takes place in the space. The condenser and evaporator comprise heat exchangers for adding and subtracting heat from the chiller refrigerant. Compressors rely on mechanical means, such as twin screws, reciprocating pistons or scrolls, to compress and expel the refrigerant at a higher pressure to push fluid through the system. From the compressor, heated and pressurized refrigerant is directed to the condenser where the refrigerant cools and condenses before reaching the expansion device. The expansion device converts the refrigerant to two-phase refrigerant comprising both liquid and gaseous components. The evaporator then vaporizes the cooled refrigerant before being returned to the compressor to continue the vapor-compression process.

In chiller systems, the evaporator is used to service a “cooler” heat exchanger that circulates a coolant such as refrigerant, or water, to cool the plurality of spaces within the building. Often, the evaporator is of a shell and tube construction to facilitate the large flow rates required of chiller systems. In such a construction, the chiller refrigerant flows through the shell to interact with a bundle of heat exchange tubes. The heat exchange tubes circulate the coolant from the cooler heat exchanger through the shell whereby the chiller refrigerant extracts heat from the coolant to supply lower temperature coolant to the cooler heat exchanger. Thus, the efficiency of the chiller and cooler systems are affected by the heat exchange between the two-phase refrigerant entering the shell and the coolant entering the tubes.

One particular problem associated with heat exchange evaporators arises from the two-phase nature of the refrigerant flowing across the heat exchange tubes. Typically, the two-phase refrigerant enters the shell at a single inlet located near the middle of a width of the shell. As the gaseous and liquid refrigerant enters the shell, the liquid refrigerant settles into a pool at the bottom of the shell, while the gaseous refrigerant rises through the liquid refrigerant to cause a cavity or bubble in the liquid pool. As such, a void is created that prevents the liquid refrigerant from exchanging heat with the heat exchange tubes, thereby interfering with the vaporization process and the efficiency of the evaporator.

Refrigerant distributors are therefore used to disperse the two-phase refrigerant across the shell such that the liquid refrigerant reaches a greater surface area of the heat exchange tubes. In addition, having the gaseous refrigerant dispersed evenly within the shell promotes convection boiling in the liquid pool from the tubes. However, difficulties arise in implementing and fabricating conventional distributors. For example, distributors often produce a large pressure drop across the distributor, which reduces the available pressure for circulating refrigerant through the vapor-compression circuit. Such pressure drops are particularly problematic for vapor-compression circuits using economizers. Also, such distributors often comprise intricate assemblies that add production time and cost to manufacturing the evaporator. There is, therefore, a need for an improved distributor that overcomes these and other problems.

SUMMARY

The present invention is directed to a distributor for an evaporator used in a vapor-compression circuit. The distributor comprises an elongate body defining an inlet portion and first and second distal ends. The body being positionable along an interior sidewall of the evaporator. The inlet portion positionable opposite an inlet port of the evaporator such that refrigerant entering the evaporator engages the inlet portion of the distributor. The first and second distal ends extending outwardly towards opposing side ends of the evaporator. The elongate body cooperates with the sidewall of the evaporator to define a channel having a generally uniform cross-sectional area extending from the first distal end to the second distal end such that refrigerant entering the channel near the inlet portion is initially substantially contained between the sidewall of the evaporator and the elongate body. The elongate body defining discharge ports to disperse refrigerant from the channel into the evaporator.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a vapor-compression system including an evaporator having a distributor of the present invention.

FIG. 2 shows a front diagrammatic view of a shell and tube heat exchanger having a distributor of the present invention.

FIG. 3 shows a cutaway perspective view of an evaporator having a distributor of the present invention:

FIG. 4 shows a front view of a first embodiment of a distributor of the present invention in which refrigerant is released along discharge ports having increasing cross sectional area along a length of the distributor.

FIG. 5 shows a front view of a second embodiment of a distributor of the present invention in which refrigerant is released along successive discharge ports having increasing heights.

FIG. 6 shows a front view of a third embodiment of a distributor of the present invention in which refrigerant is released along successive discharge ports having increasing widths.

DETAILED DESCRIPTION

FIG. 1 shows a schematic of vapor-compression system 10 including distributor 12 of the present invention. Vapor-compression system 10 includes discharge piping 14A, suction piping 14B, condenser piping 14C, evaporator piping 14D, compressor 16, condenser 18, expansion device 20 and evaporator 22. Compressor 16, condenser 18, expansion device 20 and evaporator 22 are connected in a series circuit using conduit including compressor discharge piping 14A, compressor suction piping 14B, condenser piping 14C and evaporator piping 14D. Vapor-compression system 10 also includes other components such as economizer 24 and oil distribution system 26. In one embodiment, vapor-compression system 10 comprises a water cooled “chiller” system that is used to provide cooling to a plurality of spaces, such as within a building. Condenser 18 and evaporator 22 comprise heat exchangers in which chiller refrigerant is circulated to exchange heat with fluids traveling through tube bundles 28 and 30, respectively. Condenser 18 includes manifolds 32A and 32B that conduct water from a cooling tower through tube bundle 28. The water tower cools water that is used to transfer heat from the chiller system. Evaporator 22 also includes manifolds 34A and 34B that conduct a coolant fluid, such as water, refrigerant or a blend, from a “cooler” heat exchanger through tube bundle 30. The cooler heat exchanger services one or more heat exchangers used to cool the plurality of spaces.

In the embodiment shown, the vapor-compression system 10 comprises rotary screw compressor 16 that compresses a refrigerant, such as R-122 or R-134a, to provide heated, high pressure refrigerant to condenser 18 through discharge line 14A. In other embodiments, compressor 16 includes other mechanical means for compressing a working fluid, such as reciprocating pistons or orbiting scrolls. For any mechanical compression means, compressor 16 is provided with a source of oil from oil distribution system 26 to provide cooling and lubrication to compressor 16. The oil is mixed with the refrigerant within compressor 16 and both are delivered to condenser 18 through discharge line 14A. The oil is filtered from the refrigerant within condenser 18 through an oil separator 36 that collects and returns the oil to compressor 16 with distribution system 26. Using cooling water from the cooling tower provided through manifold 32B, the refrigerant cools and condenses to a saturated liquid having a slightly lower temperature at a high pressure within condenser 18, rejecting heat to the water within tube bundle 28.

From condenser 18, the refrigerant is conducted through condenser piping 14C to expansion device 20 whereby the refrigerant undergoes a flash evaporation process to a reduced pressure and temperature and is converted to two-phase refrigerant comprising gaseous and liquid phase refrigerant. The gaseous refrigerant is vented back to compressor 16, such as through ecomonizer 24. Under pressure from compressor 16, the two-phase refrigerant continues through evaporator piping 14D to evaporator 22 at inlet port 38. In order to improve efficiency of vapor compression system 10, particularly the heat transfer efficiency of evaporator 22, distributor 12 is provided in evaporator 22. Distributor 12 is positioned within evaporator 22 to receive the two-phase refrigerant produced by expansion device 20. Distributor 12 prevents the accumulation of vapor phase refrigerant near inlet port 38 such that more surface area of tube bundle 30 is available to contact liquid phase refrigerant increasing heat transfer between the chiller refrigerant and the coolant from manifold 34B. The relative warmth of the coolant from the “cooler” heat exchanger provided by manifold 34B vaporizes the chiller refrigerant into a saturated vapor phase refrigerant. Under suction from compressor 16, the refrigerant exits evaporator 22 at outlet port 40 and returns to compressor 16 through suction piping 14B. As such, vapor-compression system 10 operates using well-known thermodynamic principles to transfer heat from evaporator 22 to condenser 18.

FIG. 2 shows a front diagrammatic view of shell and tube evaporator 22 having distributor 12 in accordance with the present invention. Evaporator 22 also includes tube bundle 30, manifolds 34A and 34B, inlet port 38, outlet port 40 and shell 42. Shell 42 comprises a flooded type evaporator in which a cylindrical body envelops tube bundle 30 and a pool of liquid refrigerant. The sides of shell 42 are capped by manifolds 34A and 34B to form a pressure vessel into which refrigerant flows to flood shell 42 and substantially immerse tube bundle 30 in the liquid refrigerant pool. Tubes of tube bundle 30 extend across shell 42 between manifolds 34A and 34B. Inlet port 38 is positioned near a bottom portion of shell 42 and outlet port 40 is positioned near a top portion of shell 42. In the illustrated embodiment, distributor 12 is positioned at the bottom of shell 42 opposite the inlet port 38.

Refrigerant flowing through vapor-compression system 10 is partially vaporized or throttled at expansion device 20 (FIG. 1) and directed into evaporator 22 at inlet port 38 as two-phase refrigerant, indicated by solid arrows in FIG. 2. Simultaneously, warmed coolant from the cooler heat exchanger is circulated through tube bundle 30 using manifold 34B, as indicated by outlined arrows in FIG. 2. The warmed refrigerant enters some of the tubes through inlet 44 at manifold 34B, flows through tube bundle 30 and into manifold 34A whereby the warmed refrigerant is turned around and directed back to manifold 34B through other tubes to leave evaporator 22 at outlet 46. In other embodiments, the warmed refrigerant enters evaporator 22 at manifold 34B, flows through all of the tubes of tube bundle 30, and leaves evaporator 22 at manifold 34A. Tube bundle 30 comprises a plurality of individual tubes between which the two-phase refrigerant passes through. The relative heat from the warmed coolant boils and vaporizes the liquid component of the two-phase chiller refrigerant through pool boiling and convective heat transfer such that vapor refrigerant leaves evaporator 22 at outlet 40.

Typically, entering gaseous refrigerant comprises about 15% to about 20% by mass of the two-phase refrigerant, which is equal to about 91% to about 95% by volume. Generally, most or all of the liquid refrigerant is vaporized by tubes 30 such that the entire volume of incoming two-phase refrigerant leaves evaporator 22 as gaseous phase refrigerant. In particular, the gaseous refrigerant has a tendency to accumulate above inlet port 38, forming a vapor barrier and inhibiting liquid refrigerant from contacting the tubes of tube bundle 30. The gaseous refrigerant forms a dry bubble around the tubes preventing boiling heat transfer from occurring. In order to increase liquid contact with tube bundle 30 and efficiency of evaporator 22, distributor 12 is positioned at the bottom of shell 42. Distributor 12 prevents the formation of large vapor bubbles by dispersing the gaseous phase refrigerant into smaller bubbles that flow through evaporator 22.

FIG. 3 shows a cutaway perspective view of evaporator 22 showing shell 42 and distributor 12 of the present invention. Shell 42 comprises a cylindrical body into which two-phase refrigerant is received at inlet port 38 and gaseous refrigerant is discharged at outlet port 40. Manifolds 34A and 34B (FIG. 2) connect with shell 42 at the open side ends of shell 42. Distributor 12 is connected to an interior surface of shell 42 such that distributor 12 covers inlet port 38. In the embodiment shown, inlet port 38 is positioned near the centers of evaporator 22 and distributor 12, but in other embodiments inlet port 38 can be positioned nearer one end of evaporator 22. Distributor 12 comprises an elongate body that forms a trough-like or tube-like structure that is open at a portion of its outer perimeter along an entire segment of its length. As such, distributor 12 cooperates with shell 42 to form channel 49 defined between shell 42, which defines a bottom wall of the channel, and distributor 12, which forms the remaining wall portions of the channel. Thus, channel 49 defines a generally uniform cross-sectional area extending substantially along the entire length of the elongate body of distributor 12. The cross-sectional profile of the elongate body of distributor 12 can define various profiles, such as rectangular, square, semi-circular or V-shaped.

In the embodiment of the invention shown in FIG. 3, distributor 12 comprises a V-shaped elongated body. The V-shaped body includes first side 50A, second side 50B, first bottom edge 52A, second bottom edge 52B and top edge 54. First side 50A and second side 50B, which are joined along top edge 54, include inlet portions 56A and 56B, distal ends 58A through 58D, and discharge ports 60A through 60D. In one embodiment, distributor 12 is fabricated from a single flat piece of steel having a perimeter that includes bottom edges 52A and 52B, inlet portions 56A and 56B and distal ends 58A through 58D. Discharge ports 60A-60D of distributor 12 are then cut into the perimeter of the flat piece. Cutting discharge ports 60A-60D at bottom edges 52A and 52B prevents fluid erosion on tube bundle 30 and simplifies production of distributor 12. The flat piece is then bent or folded to form top edge 54 and to give distributor 12 a V-shaped profile. In one embodiment, the angle between first side 50A and second side SOB is approximately 120 degrees. In other embodiments, distributor 12 is formed from a piece of stock angle iron already including the bend. In one embodiment, inlet portions 56A and 56B and distal ends 58A-58D are welded to shell 42 to form channel 49 between shell 42 and distributor 12. In other embodiment, other fastening or joining means can be used. Thus, distributor 12 can be easily manufactured from stock material with few manufacturing steps required, which helps speed production time and reduce production costs.

Referring to FIG. 3, the cross-section of channel 49 is defined by an arcuate bottom portion of an inner sidewall of the shell 42 and an opposing angular portion of the distributor 12. Thus, channel 49 extends substantially uniformly along the length of distributor 12. Inlet portions 56A and 56B are positioned opposite inlet port 38. Mass flow of two-phase refrigerant entering inlet port 38 impacts inlet portions 56A and 56B of first side 50A and second side 50B to divide into two streams of refrigerant flowing outwardly along the length of distributor 12 toward distal ends 58A-58D thereof through channel 49. The refrigerant is inhibited from migrating up toward outlet port 40 by first and second sides 50A and 50B. Distributor 12 includes discharge ports 60A-60D for controlledly dispersing the two-phase refrigerant into shell 42 such that a greater amount of liquid phase refrigerant will impact tube bundle 30 as compared to the absence of distributor 12. Thus, convective heat transfer is promoted to facilitate evaporating the liquid phase refrigerant. Furthermore, discharge ports 60A-60D are shaped to produce a minimal pressure drop in the refrigerant as it passes through distributor 12 such that pressure is retained to push the refrigerant through the vapor-compression system and to control expansion device 20. In particular, discharge ports 60A-60D are shaped to produce equal mass flow rates of the two-phase refrigerant at successive positions along first and second sides 50A and 50B. The distal ends of the V-shaped body between first side 50A and second side 50B are open to also permit two-phase refrigerant to escape from within channel 49.

In embodiment shown, discharge ports 60A-60D comprise wedge-shaped vents that are positioned along bottom edges 52A and 52B. For example, first port 60A extends along bottom edge 52A. First port 60A begins at inlet portion 56 and extends toward distal end 58A, extending increasingly into bottom edge 52A and first side 50A to form a wedge-shaped vent. The rise and run of the wedges are selected, in conjunction with the size of the cross sectional area of flow channel 49 within distributor 12, to release equal amounts of the two-phase refrigerant at different positions along distributor 12. The specific geometry of the discharge ports are selected using mathematical or numerical analysis of two-phase fluid flow to model refrigerant flow through channel 49 and discharge ports 60A-60D. For the purposes of simplification, discussion of the benefits of the present invention can be seen with reference to mass and momentum conservation equations and steady-state fluid flow.

The volumetric flow rate Q of a fluid entering a channel is determined by the area A₁ of the inlet of the channel multiplied by the velocity V₁ of the fluid at the inlet, as shown in Equation (1). The volumetric flow rate of a fluid entering the channel must equal the volumetric flow rate of the fluid leaving the channel at an outlet area A₁ and outlet velocity V₂, as is also shown in Equation (1).

Q=A₁V₁=A₂V₂ (m³/s)  Equation (1)

Thus, the volumetric flow rate of two-phase refrigerant entering channel 49 through inlet port 38 is equal to the total amount of refrigerant leaving channel 49 through discharge ports 60A-60DB and the open ends of channel 49.

The mass flow rate rh of a fluid is determined by the density of the fluid p multiplied by the volumetric flow rate Q of the fluid, as shown by Equation (2). Additionally, the mass flow rate m of the fluid can be determined based on pressures losses between two points of the flow path, P₁ & P₂, and the area of the flow path A₂, as is also shown in Equation (2), where K represents a constant based on friction and other factors.

{dot over (m)}=ρQ=ρA ₂ V ₂ =KA ₂√{square root over (2ρ(P ₁ −P ₂))} (k_(g)/s)  Equation (2)

It can be seen that mass flow rate depends on the velocity of the fluid and the area through which it is flowing at any given point. Thus, mass flow rate of a refrigerant through a discharge port depends on the area of the discharge port and the velocity of the refrigerant through the port. Alternatively, the mass flow rate through the discharge port depends on the area of the discharge port and the pressures on either side of the port. Thus, the mass flow rate of a refrigerant through a discharge port depends on the pressure in channel 49 and the pressure in shell 42, as well as the area of the discharge port and channel 49.

As is known from Bernoulli's equation and viscous head losses, the velocity and pressure of a fluid flowing through a channel decreases along the length of that channel due to friction and other considerations. Thus, for example, by manipulating the size of discharge port 60A and knowing the velocity profile of the refrigerant through channel 49, distributor 12 of the present invention produces a mass flow rate emitting from a portion of discharge port 60A near inlet portion 56A that is equal or near equal to a mass flow rate emitting from a portion of discharge port 60A near distal end 58B to produce more uniform flow of two-phase refrigerant across the width of shell 42. Thus, discharge ports having various geometries can be produced along the length of distributor 12 to achieve more uniform refrigerant distribution within evaporator 22.

FIG. 4 shows a front view of the first embodiment of distributor 12 of the present invention shown in FIGS. 2 and 3, in which refrigerant is released along discharge ports 60A and 60B. Distributor 12 includes first side 50A, bottom edge 52A, top edge 54, inlet portion 56A, distal ends 58A and 58B, discharge ports 60A and 60B and mounting posts 62A and 62B. Mounting posts 62A and 62B are configured to be received into mating bores or holes within shell 42 so that bottom edge 52A along inlet portion 56A and distal ends 58A and 58B contact the interior surface of shell 42. Posts 62A and 62B are then secured to shell 42 using a welding process or some other method. In other embodiments, posts 62A and 62B comprise appendages that are connected to bottom edge 52A. Second side 50B of distributor 12 includes similar components and features as first side 50A, including discharge ports and mounting posts. As such, in the embodiment shown in FIG. 4, distributor 12 includes four mounting posts and four wedge-shaped discharge ports.

Mass flow {dot over (m)}₁ of refrigerant enters distributor 12 at pressure P₁ and velocity V₁ near inlet portion 56A and is divided into two branches {dot over (m)}₂ and {dot over (m)}₃ each containing half the mass flow of the amount that enters at inlet port 38 (FIG. 3). The two branches {dot over (m)}₂ and {dot over (m)}₃ of two-phase refrigerant continue to distal ends 58A and 58B, losing refrigerant through discharge ports 60A and 60B. Pressure P₁ and pressure P′ within shell 42 are determined by the operating conditions of vapor-compression system 10 and the pressure loss across distributor 12. Pressure P₁ is generally greater than pressure P′ as some pressure differential across distributor 12 must exist to enable flow through distributor 12. Discharge ports 60A and 60B are shaped to produce uniform mass flow along the width w of each discharge port by varying the pressure differential between the pressure within distributor 12 near the various positions along each discharge port and P′. For example, at point X near inlet portion 56A, the refrigerant has velocity V₄ at pressure P₄ and the area of discharge port 60A near that point has a given area. Pressure P₄ is less than pressure P₁ as some head loss occurs between inlet port 38 and discharge port 60A. As refrigerant flows through distributor 12 along discharge port 60A, such as to point Y, the pressure and velocity of the refrigerant continues to decrease to pressure P₅ and velocity V₅. Thus, P₁>P₄>P₅>P′, and V₁>V₄>V₅, as would be expected from a flow channel having a nearly constant cross sectional area. In order to keep the mass flow rate along the length of distributor 12 equal, the area of discharge port 60A varies such that the area of discharge port 60A near point Y is greater than the area of discharge port 60A near point X.

As illustrated by Equation (2), in order to keep a constant mass flow rate for a flow of a fluid having a decreasing velocity, the area must increase. As Equation (2) also illustrates, in order to keep a constant mass flow rate for a flow of fluid having a decreasing pressure differential, the area must increase. The sizes of discharge port 60A, as defined by height h and width w, and the size of the channel defined by distributor 12 and shell 42, is selected such that the mass flow rates at various points along discharge port 60A, such as {dot over (m)}₄ and {dot over (m)}₃ at points X and Y are equal. The total mass flow rate leaving discharge ports 60A and 60D (FIG. 3) equals {dot over (m)}₂. In one embodiment of the invention, the length of distributor 12 can be sized such that the mass flow rates at the openings of distributor 12 near distal ends 58A and 58B is zero or nearly and the pressures there approximately equal pressure P′. The size of the discharge ports is also selected to produce the smallest pressure drops possible across the discharge ports. Generally, discharge port 60A has increasing cross sectional area along a length of distributor 12 extending from inlet portion 56A to distal end 58A. In other words, the height h of discharge port 60A increases along width w, from inlet portion 56A to distal ends 58A.

Mathematical and numerical modeling can be used to analyze the mass flow of the refrigerant at any point along the length of the discharge ports to determine distribution of the refrigerant along the length of the discharge ports. The modeling can also be used to determine and verify other considerations. For example, the modeling can be used to ensure that the geometry of channel 49 and the discharge ports is selected to produce minimal pressure loss across distributor 12. Additionally, the geometry of the distributor 12 is selected to avoid problems associated with sound velocity. In high velocity flow of refrigerant through a channel, the velocity of the refrigerant is limited by the speed of sound and the area of the channel. Distributor 12 of the present invention is sized to avoid sound velocity issues by sizing channel 49 to be large enough such that refrigerant flowing toward distal ends 58A and 58B is below the sound velocity.

The following equations demonstrate the one modeling approach for determining two-phase flow through a distributor.

${dp}_{a} = {\frac{1}{2}f\; \rho_{a}{v_{a}^{2} \cdot \frac{{dL}_{a}}{D}}}$ ${dp}_{b} = {\frac{1}{2}f\; \rho_{b}{v_{b}^{2} \cdot \frac{{dL}_{b}}{D}}}$ ${d{\overset{.}{m}}_{a}} = {\rho_{a}v_{a}{\frac{d\left( {A_{a} - A_{c}} \right)}{{dL}_{a}} \cdot {dL}_{a}}}$ ${d{\overset{.}{m}}_{b}} = {\rho_{b}v_{b}{\frac{d\left( {A_{b} - A_{c}} \right)}{{dL}_{b}} \cdot {dL}_{b}}}$

Boundary conditions for pressure and mass flow balance are:

Δp_(a)=Δp_(b)

{dot over (m)}+{dot over (m)}={dot over (m)}

That is,

∫₀^(L_(a))p_(a) = ∫₀^(L_(b))p_(b) ${{\overset{.}{m}}_{a,e} + {\overset{.}{m}}_{b,e} + {\int_{0}^{L_{a}}{{\overset{.}{m}}_{a}}} + {\int_{0}^{L_{b}}{{\overset{.}{m}}_{b}}}} = \overset{.}{m}$

Where,

{dot over (m)}_(a,e)=ρ_(a,e)ν_(a,e)A_(c)

{dot over (m)}_(b,e)=ρ_(b,e)ν_(b,e)A_(c)

f Darcy friction factor using Martinelli two-phase model

A_(a),A_(b) Open flow area at each section

A_(c) Cross section area of the distributor

D Hydraulic diameter of the cross section of distributor

e Subscription indicating the end of distributor

L_(a) Length of one half of distributor from inlet portion to distal portion

L_(b) Length of one half of distributor from inlet portion to distal portion

These equations illustrate, among other things, that flow entering the distributor equals the flow exiting the distributor through the discharge ports. The flow is determined by, among other things, integrating the pressure differential along the length of each discharge port. Thus, mathematical and numerical modeling can be used to determine the cross sectional area of channel 49 within distributor 12 and the shape of discharge ports 60A-60D such that mass flow discharged at successive points along the width of distributor 12 are equal. Using discharge channels 60A-60D, distributor 12 of the present invention achieves true uniform distribution of refrigerant mass flow across the width of each discharge port. In other embodiments of the invention, a series of discharge ports is used to produce successive distributions of two-phase refrigerant having the same mass flow rate, as are shown in FIGS. 5 and 6.

FIG. 5 shows a front view of a second embodiment of distributor 12 of the present invention in which refrigerant is released along successive, discrete discharge ports having increasing opening heights. FIG. 6, which is discussed concurrently with FIG. 5, shows a front view of a third embodiment of a distributor of the present invention in which refrigerant is released along successive discharge ports having increasing opening widths. The distributors of FIGS. 5 and 6 are fabricated from similar processes and materials. However, rather than shaping bottom edge 52A to form the discharge ports, the discharge ports are cut into sides 50A and 50B (FIG. 3), such as with a laser cutting process. As such, distributor 12 comprises a rectangular piece of steel that is bent to provide top edge 54, and bottom edges 52A and 52B (FIG. 3) contact the inner surface of shell 42 along their entire length. In the embodiments of FIGS. 5 and 6, wedge-shaped discharge ports 60A and 60B have each been replaced with a pair of quadrangular discharge ports. For example, first side 50A of distributor 12 includes discharge ports 64A-64D. Second side 50B of distributor 12 also includes four quadrangular discharge ports such that, in the embodiment shown, distributor 12 includes eight discharge ports. However, in other embodiments other numbers of discharge ports can be used.

The embodiments of distributor 12 shown in FIGS. 5 and 6 operate similar to that of distributor 12 in FIG. 4. However, rather than producing a continuous release of two-phase refrigerant along the length of the distributor, intermittent bursts of refrigerant are released at discrete points along the distributor, with a greater number of ports producing more uniform mass distribution. Mass flow {dot over (m)}₁ of refrigerant enters distributor 12 at pressure P₁ and velocity V₁ near inlet portion 56A and is divided into two branches {dot over (m)}₂ and {dot over (m)}₃ each containing half the mass flow of the amount that enters at inlet port 38. The velocity of mass flow {dot over (m)}₂ decreases as it travels past discharge port 64A and 64B towards distal end 58A, losing pressure along the way due to head losses. In order to produce uniform mass distribution from each discharge port, the areas of the discharge ports are increased as the series of discharge ports extends toward distal end 58A.

In FIG. 5, discharge port 64A comprises a rectangle having height h₁ and width w₁, and discharge port 64B comprises a rectangle having height h₂ and width w₂. The widths w₁ and w₂ are equal, but height h₂ is greater than height h₁ such that discharge port 64B has a greater area than discharge port 64B. In FIG. 6, discharge port 64A comprises a rectangle having height h₁ and width w₁, and discharge port 64B comprises a rectangle having height h₂ and width w₂. Widths W₁ is greater than width w₂, and height h₁ is equal to height h₂ such that discharge port 64B has a greater area than discharge port 64B. Discharge ports 64A and 64B are bounded by interior edges within side 50A. Additionally, in other embodiments, other shaped discharge ports, such as circular, can be used. In another embodiment, triangular discharge port having a shape similar to the wedge-shaped discharge ports of FIG. 4 can be incorporated into the perimeter of distributor 12 to produce more uniform distribution of the refrigerant. The difference in pressure between P′ and the pressure at discharge port 64A is greater than the difference in pressure between P′ and the pressure at discharge port 64B. As such, the mass flow rates {dot over (m)}₄ and {dot over (m)}₃ are equal.

Distributor 12 of the present invention produces homogeneous distribution of two-phase refrigerant within an evaporator to optimize evaporation of liquid phase refrigerant. Distributor 12 is produced from a simple, single-piece body that can be fabricated from a single, flat piece of stock steel material, or from a stock piece of angle iron. Distributor 12 includes discharge ports that produce equal mass flows of refrigerant at various positions along the width of distributor 12. In one embodiment, the discharge ports comprise wedge-shaped slots that release a continuum of refrigerant to produce true uniform distribution. In another embodiment, the discharge ports comprise successively larger windows in distributor 12 that release discrete and equal mass flows of refrigerant. Computer and numerical modeling can be used to optimize the flow, minimize pressure losses and to account for sound velocity. The geometry of distributor 12 and the discharge ports is easily scaled up or down for use in evaporators having different capacities.

Although the present invention has been described with reference to preferred embodiments, workers skilled in the art will recognize that changes may be made in form and detail without departing from the spirit and scope of the invention. 

1. A distributor for use in an evaporator for a vapor-compression circuit, the distributor comprising: an elongate body positionable within the evaporator adjacent a sidewall thereof; an inlet portion positionable near an inlet port of the evaporator to receive a flow of refrigerant; first and second distal ends extending along a length of the evaporator; discharge ports positioned along a length of the elongate body; and a channel having a cross-sectional area extending from the first distal end to the second distal end, wherein the channel is defined by the elongate body and the sidewall of the evaporator.
 2. The distributor of claim 1 wherein the elongate body comprises: first and second side surfaces for including the discharge ports; a bottom edge disposed along each side surface for contacting the sidewall of the evaporator; and a top edge along which the first and second side surfaces are joined.
 3. The distributor of claim 2 wherein the channel has a V-shaped cross section such that the channel has a substantially constant cross-sectional area.
 4. The distributor of claim 2 wherein the discharge ports comprise elongate, tapered openings having increasing area extending from near the inlet portion to near the first and second distal ends, respectively.
 5. The distributor of claim 4 wherein the discharge ports comprise cut away portions of the side surfaces extending into the bottom edge.
 6. The distributor of claim 4 wherein the discharge ports are configured to be bounded by the bottom edge of the elongate body and the sidewall of the evaporator.
 7. The distributor of claim 2 wherein the discharge ports comprise a series of windows positioned in the side surfaces that extend from near the inlet portion to near the first and second distal ends, respectively.
 8. The distributor of claim 7 wherein the windows are bounded by interior edges within the side surfaces.
 9. The distributor of claim 7 wherein the series of windows comprise rectangles extending into the side surfaces such that successive windows have larger areas.
 10. The distributor of claim 7 wherein the series of windows comprise rectangles extending into the side surfaces such that successive windows have larger widths.
 11. The distributor of claim 7 wherein the series of windows comprise rectangles extending into the side surfaces such that successive windows have larger heights.
 12. The distributor of claim 1 wherein the first and second distal ends of the elongate body are open to permit refrigerant to escape.
 13. The distributor of claim 1 wherein the discharge ports are configured to discharge an equal amount of mass of refrigerant at various positions along the elongate body.
 14. The distributor of claim 13 wherein the cross section of the channel is configured to produce a decrease in velocity of refrigerant flowing from the inlet portion to the first and second distal ends, and the size of the discharge ports are configured to produce decreasing pressure differentials along the elongate body from the inlet portion to the first and second distal ends.
 15. An evaporator for use in a vapor-compression circuit, the evaporator comprising: an annular shell body comprising: an inlet port positioned on a wall of the annular shell body; an outlet port positioned across from the inlet port; a tube bundle extending generally across an interior length of the annular shell body; a distributor casing extending across the an interior of the annular shell body and covering the inlet port such that a channel having a cross section is formed by the distributor casing and the annular shell body, the distributor casing comprising discharge ports positioned along a length of the distributor casing.
 16. The evaporator of claim 15 wherein the discharge ports are configured to release an equal amount of mass of refrigerant at various positions along the channel.
 17. The evaporator of claim 15 wherein the cross section of the channel is configured to produce a decrease in velocity of refrigerant flowing through the channel from near the inlet port, and the size of the discharge ports are configured to produce decreasing pressure differentials across the discharge ports extending from near the inlet and along the channel.
 18. The evaporator of claim 15 wherein the discharge ports comprise openings having increasingly larger areas extending from near the inlet port through the channel.
 19. A method for distributing two-phase refrigerant in an evaporator using a distributor, the method comprising: introducing refrigerant into an evaporator; channeling the refrigerant into opposing legs of a distributor, the legs having substantially constant cross-sectional areas; disbursing the refrigerant from the channel into the evaporator through discharge ports positioned along lengths of the legs, the discharge ports having increasing areas extending along the lengths of the legs.
 20. The method of claim 20 wherein the step of disbursing the refrigerant further comprises continuously releasing refrigerant having uniform mass flow along lengths of the discharge ports. 